Controlling cylinder mixture and turbocharger operation

ABSTRACT

A method for controlling differences in exhaust gas residual amount for a two cylinder bank engine having at least one turbocharger is presented. In one example, the description includes a method for adjusting valve timing to reduce cylinder exhaust gas residual variation.

CROSS REFERENCE TO RELATED APPLICATIONS

The present application is a continuation of U.S. patent applicationSer. No. 13/168,584, filed Jun. 24, 2011, now U.S. Pat. No. 8,180,553,which is a continuation of U.S. patent application Ser. No. 12/882,077,filed Sep. 14, 2010, now U.S. Pat. No. 7,987,040, which is acontinuation of U.S. patent application Ser. No. 11/777,591, filed Jul.13, 2007 and now U.S. Pat. No. 7,801,665, the entire contents of each ofwhich are incorporated herein by reference.

FIELD

The present description relates to a method for controlling a cylindermixture of an internal combustion engine having a turbocharger and avariable event valvetrain.

BACKGROUND

One method to operate an internal combustion engine having aturbocharger is presented in U.S. Pat. No. 6,202,414. This methodpresents a way to balance the output of two turbochargers that are incommunication with different cylinder banks of a “V” engine. Balancingthe output of two turbochargers can improve turbocharger efficiency andmitigate the possibility of compressor surge. This method uses stateparameters that are in the vicinity of the turbocharger compressors toadjust the effective turbine cross section, thereby regulatingcompressor flow. The method attempts to reduce any difference in stateparameters between two cylinder groups to zero. That is, the methodattempts to equalize the state parameters of the two cylinder groupsthat are in communication with the two turbochargers.

The above-mentioned method can also have several disadvantages. Forexample, the method simply adjusts a turbocharger waste gate or variablegeometry vane position, thereby simultaneously affecting turbochargerand cylinder operation. Specifically, turbine energy is altered alongwith the amount of air and exhaust gas residuals that comprise thecylinder mixture. In other words, adjusting the waste gate or vaneposition changes the exhaust backpressure and can affect the amount ofexhaust gas residuals that are trapped in a cylinder since the exhaustbackpressure acts to impede flow from the cylinder. Further, whencylinder volume is displaced by additional residual exhaust gas, thereis less available space for fresh air in the cylinder. Consequently,cylinder charge air amount and cylinder charge residual amount maychange in an undesirable manner when a waste gate or turbine vanes areadjusted to regulate a turbocharger. As a result, regulation and/orcontrol of a turbocharger may have undesirable effects on engineemissions, fuel consumption, performance, audible noise, and uniformcylinder torque production.

The inventors herein have recognized the above-mentioned disadvantagesand have developed a method to control an engine having variable eventvalve operation and at least one turbocharger.

SUMMARY

One embodiment of the present description includes a method forcontrolling an internal combustion engine, the method comprising:varying valve operation of at least a cylinder in a second group ofcylinders as a waste gate position of a first turbocharger is adjusted,said first turbocharger located in the exhaust path of a first group ofcylinders. This method overcomes at least some of the limitations of thepreviously mentioned methods.

Variable valve operation provides an additional degree of freedom toequalize cylinder mixtures and balance mass flow through differentcylinder groups of a turbocharged engine. By adjusting valve timingrelative to crankshaft position and/or by adjusting valve lift, mixturedifferences and flow differences between different cylinder groups of aturbocharged engine can be reduced. For example, a “V” engine having twocylinder groups and two turbochargers can have less cylinder mixturevariation between cylinder banks when engine valve timing is used toequalize exhaust gas residuals between the cylinder banks. Specifically,valve timing can be used to adjust the exhaust gas residual amount andcylinder air charge at a given intake manifold pressure. Therefore, ifone bank of a turbocharged cylinder experiences a different cylinderbackpressure than the other cylinder bank, due to different waste gatepositions of the two turbochargers for example, valve timing adjustmentscan be used to mitigate differences in exhaust gas residuals andcylinder charge mixtures between the cylinder banks.

In another aspect of the present description, engine valve adjustmentscan be made in conjunction with turbocharger waste gate or vaneadjustments to equalize flow between compressors while also reducingvariation of exhaust gas residuals between cylinder banks. In oneexample, the position of turbocharger waste gates and the valve timingof exhaust valves can be adjusted to regulate energy supplied to theturbines and exhaust gas residuals of an engine having two cylinderbanks. If a compressor approaches a surge condition, its waste gate canbe opened and exhaust valve timing retarded so that residuals aremaintained at the same time that the amount of exhaust gas energydelivered to the turbocharger is lowered.

The present description may provide several advantages over priorsystems and methods. Specifically, the present method can achievedesired cylinder mixtures while regulating turbocharger operation, atleast during some conditions. Further, the method can improve engineemissions because cylinder exhaust gas residuals can be controlled evenif there are exhaust backpressure differences between cylinder banks. Inaddition, the method can produce more uniform cylinder mixtures andprovide torque that is more uniform or equal between engine cylinderbanks or cylinders.

The present method also provides unexpected results. Specifically, themethod makes it possible to make finer adjustments to the turbochargercompressor flow so that the compressor may be operated closer to thecompressor surge region. This can extend the turbocharger operatingenvelope. Also, the method can improve turbocharger durability andreduce warranty because the possibility of entering surge conditionswhen regulating compressor flow at low flow conditions is reduced.

The above advantages and other advantages and features of the presentdescription will be readily apparent from the following DetailedDescription when taken alone or in connection with the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The advantages described herein will be more fully understood by readingan example of an embodiment, referred to herein as the DetailedDescription, when taken alone or with reference to the drawings,wherein:

FIG. 1 is a schematic diagram of a turbocharged engine;

FIG. 2A is a flow chart of an example engine control strategy;

FIG. 2B is flow chart of another example engine control strategy;

FIG. 3 is a plot that shows example compressor flow characteristics;

FIG. 4 is a plot that shows the relationship between intake and exhaustvalve timing on engine volumetric efficiency;

FIG. 5 is a plot of the effect of variable exhaust valve timing oncylinder airflow; and

FIG. 6 is a plot of the effect of variable intake valve timing oncylinder airflow.

DETAILED DESCRIPTION

Referring to FIG. 1, internal combustion engine 10 is controlled byelectronic engine controller 12. Engine 10 includes a plurality ofcylinders in a “V” configuration that are similar to cylinder 50.Cylinder banks 13 and 14 are comprised of three cylinders each. Intakecams 21 and 22 operate intake valves (not shown) to regulate airflowinto the cylinders of banks 13 and 14. Exhaust cams 20 and 23 operateexhaust valves (not shown) to regulate exhaust flow out of cylinderbanks 13 and 14. Timing of intake and exhaust cams relative tocrankshaft position can be varied by adjusting phasors 76, 74, 72, and70. Alternatively, one or more intake or exhaust valves may be operatedwithout assistance from a mechanical cam (e.g., electrically orhydraulically actuated valves). Further, intake and/or exhaust valvesmay be configured to vary valve lift. Each cylinder surrounds a pistonthat transfers combustion energy to mechanical energy through crankshaft40. Intake manifold 44 is in communication with electronicallycontrolled throttle 125 and directs air to cylinder banks 13 and 14.Intake air is routed through duct 42 and mass airflow sensor 60 beforebeing compressed by first turbocharger 30. A second turbocharger 34,also compresses air from duct 42. Compressed air is routed through duct43 to intercooler 50 and proceeds to the inlet of electronic throttle125. Combusted gases exit cylinder banks 13 and 14 through exhaustmanifolds 52 and 54. Exhaust gases rotate turbines 31 and 36 ofturbochargers 30 and 34, turbines 31 and 36 cause compressors 32 and 35to rotate compressing fresh air. Waste gates 33 and 37 (i.e.,turbocharger control actuators) allow exhaust gases to bypass turbines31 and 36 so that turbine flow of each turbocharger can be controlled.Alternatively, turbines 31 and 35 may have adjustable vanes (i.e.,turbocharger control actuators) that allow the compressor and turbineflow to be controlled. Also, if desired, a single turbocharger can beused to compress fresh air for both cylinder banks.

Distributorless ignition system (not shown) provides ignition spark tothe cylinders of banks 13 and 14. Universal Exhaust Gas Oxygen (UEGO)sensors 85 and 86 are shown coupled to exhaust manifolds 52 and 54upstream of catalytic converters 70 and 71. Alternatively, two-stateexhaust gas oxygen sensors may be substituted for UEGO sensors 85 and86. Two-state exhaust gas oxygen sensor 98 is shown coupled to exhaustpipe 49 downstream of catalytic converter 70. Alternatively, sensor 98can also be a UEGO sensor. A second two-state oxygen sensor 99 is shownsimilarly situated. Catalytic converter temperature is measured bytemperature sensor 77, and/or estimated based on operating conditionssuch as engine speed, load, air temperature, engine temperature, and/orairflow, or combinations thereof.

Converter 70 can include multiple catalyst bricks, in one example. Inanother example, multiple emission control devices, each with multiplebricks, can be used. Converter 70 can be a three-way type catalyst inone example. A second catalytic converter 71 processes exhaust gases onthe opposite cylinder bank.

Controller 12 is shown in FIG. 1 as a conventional microcomputerincluding: microprocessor unit 102, input/output ports 104, andread-only-memory 106, random-access-memory 108, 110 Keep-alive-memory,and a conventional data bus. Controller 12 is shown receiving varioussignals from sensors coupled to engine 10, in addition to those signalspreviously discussed, including: engine coolant temperature (ECT) fromtemperature sensor 112; a position sensor 119 coupled to an acceleratorpedal; a measurement of engine manifold pressure (MAP) from pressuresensor 122 coupled to intake manifold 44; crankshaft position 118;camshaft positions 110-113; exhaust manifold pressure 62-63; throttleinlet air pressure 115; throttle inlet air temperature 117, and ameasurement (ACT) of engine air amount temperature or manifoldtemperature from temperature sensor 124.

The engine described in FIG. 1 may be the primary means of generatingmotive force in a vehicle or it may be part of a vehicle having morethan one means for generating motive force, a hybrid vehicle forexample. The engine may generate wheel torque in conjunction with anelectric motor when in a hybrid configuration. Alternatively, the enginemay generate wheel torque in conjunction with a hydraulic motor. Thus,there are many configurations whereby features of the presentdescription may be used to advantage.

Referring now to FIG. 2, a flow chart of a method for simultaneouslycontrolling flow through two turbochargers is shown.

Engines can be configured with two turbochargers to reduce turbochargerlag and the lack of engine torque response associated with it.Turbocharger lag is caused by the time it takes air to pass through anengine, combust with fuel, and deliver energy to a turbine. Theturbine's inertial resistance to motion also contributes to lag. Byinstalling two low inertia turbochargers in a two cylinder bank engine,the turbochargers can spin up faster and reach efficient flow conditionsso that compressed air is delivered to the engine cylinders faster.However, lower inertia turbochargers can have lower pumping capacitiesthan higher inertial turbochargers. Consequently, two low inertiaturbochargers can be used in different cylinder banks to achieve pumpingcapacity that is similar to that of a larger turbocharger whileretaining lower inertia. Thus, torque response can be improved in a twinturbocharged system without reducing high load performance. However,operating two turbochargers presents a challenge of ensuringsubstantially equal performance between the two compressors so that thecompressors do not enter surge conditions. The method of FIG. 2 can beused to balance compressor performance.

In step 201, the routine determines engine operating conditions. Engineoperating conditions can be determined and/or monitored by sampling datafrom various vehicle and/or engine sensors shown, but not limited tothose illustrated in FIG. 1. For example, engine temperature can bedetermined by sampling the output of the engine coolant temperaturesensor 112 that is shown in FIG. 1. In addition, engine operatingconditions can include, but are not limited to the pressure in theengine intake manifold, the air temperature in the engine intakemanifold, the pressure in the engine exhaust manifolds, the temperaturein the engine exhaust manifolds, the state of an engine exhaust gasoxygen sensors, the engine torque request, engine speed, and barometricpressure. After engine operating conditions are determined, the routineproceeds to step 203.

In step 203, the desired air mass flow rate is determined from thecurrent engine speed and an operator torque request. The operator torquerequest may be determined from observing the position of an acceleratorpedal, a lever, or from a signal produced by an external system (e.g.,an analog or digital command from a hybrid vehicle controller). Thetorque request signal is transformed into a driver brake torque requestby way of a two dimensional table that contains engine speed data versesaccelerator pedal position. When queried, the table outputs a desireddriver brake torque. The desired driver brake torque is summed with thetorque used to operate accessories and an estimate of engine pumpingtorque to calculate a desired engine brake torque. This is illustratedby the following equation:Γ_(eng) _(—) _(brake) _(—) _(tor)=Γ_(driver)+Γ_(acc)+Γ_(pump)where Γ_(eng) _(—) _(brake) _(—) _(tor) is the requested engine braketorque, Γ_(driver) is the requested or desired driver torque demand,Γ_(acc) is the torque to power accessories (e.g., an air conditioningcompressor), and Γ_(pump) is the engine pumping torque. The requestedengine brake torque and current engine speed are then used to index atable that outputs an engine load. The engine load is converted to anair mass flow rate by the following equation:

${des\_ am} = {{sarchg}*N*\frac{Num\_ cyl}{2}*{Load}}$where des_am is the desired air mass flow rate of the engine, sarchg isthe cylinder air charge capacity based on the cylinder volume atstandard temperature and pressure, N is the engine speed, Num_cyl is thenumber of engine cylinders, and Load is a fractional number thatrepresents an engine air capacity. The specific air charge in a cylindercan be determined by the equation:Cyl_air_chg=Load*sarchg

The desired cylinder air charge can then be converted into an intakemanifold pressure by using the relationship between cylinder air chargeand intake manifold pressure. Specifically, in some variable camconfigurations, a nearly linear relationship exists between cylinder aircharge and intake manifold pressure, see FIGS. 5-6 for example, at agiven engine speed and cam positions. A family of curves that are basedon engine speed and cam positions can be stored that describe therelationship between cam positions, engine speed, cylinder air charge,and manifold pressure. These tables can be indexed by using the currentengine speed, cam positions, and desired cylinder air charge todetermine a desired intake manifold pressure. The desired intakemanifold pressure is determined by interpolating between a limitednumber of empirical curves that describe cylinder air charge in relationto cam positions and intake manifold pressure. In other variable camconfigurations, a quadratic relationship exists between intake manifoldpressure and cylinder airflow. In these cases, it is possible to solvethe roots of the equation and obtain an equation that outputs manifoldpressure as a function of cylinder air charge at given cam positions.U.S. patent application Ser. No. 11/423,433 describes the relationshipbetween cylinder air charge and manifold pressure and is hereby fullyincorporated by reference for all intents and purposes.

The desired manifold pressure is achieved by setting the position of thethrottle plate to match the desired cylinder airflow. Specifically, thethrottle position is set based on the desired pressure ratio across thethrottle (i.e., the pressure ratio between desired manifold pressure andthrottle inlet pressure) and a throttle angle that produces the desiredflow at the desired pressure ratio across the throttle. If desired, thethrottle plate position can be adjusted further using aproportional/integral controller that moves the throttle plate based onfeedback from a manifold pressure sensor.

For applications having valves that can be operated independent ofcrankshaft position, electrically actuated valves for example, themethod described in U.S. Pat. No. 7,079,935, which is hereby fullyincorporated by reference for all purposes, may be used to determinevalve timing. In this embodiment, the desired manifold pressure can bedetermined by indexing a table that outputs desired manifold pressureusing engine speed and operator torque demand. Throttle position can beset as described above. The routine proceeds to step 205.

In step 205, the turbocharger waste gate position or vane position isset. Engine speed and desired engine load are used to index a table thatoutputs a desired waste gate position. The waste gate position can befurther adjusted for a throttle inlet pressure error (i.e., the desiredthrottle inlet pressure minus the actual throttle inlet, pressure) byadjusting the waste gate position in response to the throttle inletpressure error. In one example, a proportional/integral controller canbe used to make waste gate adjustments based on the throttle inletpressure error. The routine continues to step 207.

In step 207, the open-loop camshaft positions or valve timings aredetermined. Alternatively, valve lifts may be determined in this step insubstantially the same manner, i.e., by indexing predetermined tablesand/or functions. Engine speed and desired engine torque are used toindex tables that output empirically predetermined cam positions orvalve timings for each cylinder bank. The routine proceeds to step 209.

In step 209, the routine determines whether or not to make valve timing,valve lift, and/or waste gate/vane adjustments to correct for flowdifferences between two compressors. For example, when an engine isoperating at 1500 RPM, the mass flow of one compressor can exceed themass flow rate through the other compressor. As such, the valvetiming/lift of one or both cylinder banks can be adjusted to reducevariations in compressor flow rates.

To determine whether or not valve adjustments are desirable, acomparison is made between the mass of air flowing through eachcompressor. In one example where a single mass airflow sensor is used,it is possible to measure flow through one compressor and estimate theflow through the other compressor. The estimated compressor flow can bedetermined from the following equation:{dot over (m)} _(c,2)=am−{dot over (m)} _(c,1)where am is the engine air mass which can be determined from enginespeed, intake manifold pressure, volumetric efficiency, and the idealgas law PV=nRT, {dot over (m)}_(c,1) is the measured flow rate throughcompressor 32, and {dot over (m)}_(c,2) is the flow rate throughcompressor 35. To improve the estimate during transient conditions thefollowing equation may be used:

${\frac{s}{{\tau\; s} + 1}P_{c,{out}}} = {\frac{{RT}_{tct}}{V_{tim}}\left( \frac{1}{{\tau\; s} + 1} \right)\left( {{\overset{.}{m}}_{c,1} + {\overset{.}{m}}_{c,2} - {\overset{.}{m}}_{thr}} \right)}$This expression can be solved for {dot over (m)}_(c,2), with s being theLaplace variable and

$\frac{s}{{\tau\; s} + 1}\mspace{14mu}{and}\mspace{14mu}\frac{1}{{\tau\; s} + 1}$denoting a nigh pass filter and a low pass filter respectively.P_(c,out) is defined as the pressure at the compressor outlet, which canbe estimated from measured pressure at the throttle inlet, P_(tip), andthe estimated pressure loss across the intercooler; T_(tct) is themeasured temperature at the throttle inlet; and {dot over (m)}_(thr) isthe mass flow rate through the throttle, which can be determined asdescribed above.

Once the individual compressor flow rates have been determined, thecompressor flow rates are compared to determine if an imbalance exists.In one example, the comparison is made using the following equation:comp_error={dot over (m)} _(c,1) −{dot over (m)} _(c,2)where comp_error represents the compressor flow error between the twoturbochargers, {dot over (m)}_(c,1) is compressor number one flow rate,and {dot over (m)}_(c,2) is compressor number two flow rate. After thecompressor flow error is determined, the routine can determine if theerror is large enough to warrant a correction of valve timing/lift. Ifdesired, the routine can limit compressor flow corrections to one ormore predetermined magnitudes of error. That is, if the compressor flowerror is below some predetermined value, then no correction will be madeto compensate for the error. In another embodiment, the flow errorbetween compressors can be compensated as long as the error is present.

In step 209, the routine also determines if adjustments to valvetiming/lift, throttle plate position, and turbocharger (e.g., waste gateor turbine vane position adjustments) are to be made to limit thepossibility of compressor surge. If the compressor is operating near orgoes into the compressor surge boundary, the valve timing/lift can beadjusted such that the engine's volumetric efficiency is decreased,thereby lowering the pressure ratio across the compressor and reducingthe possibility of compressor surge. In this example the ignition sparkmay be advanced during some conditions to offset a portion of any torqueloss that may occur as a result of the valve timing adjustment. As aresult, the possibility of compressor surge can be mitigated byadjusting valve timing/lift with less variation in engine torque.

On the other hand, consider a condition where it is desirable toincrease flow through at least one compressor. When, for example, bestvolumetric efficiency occurs when both IVO and EVO occur attop-dead-center (TDC), a cam timing that opens intake valves after TDCcan be advanced toward TDC so that engine volumetric efficiencyincreases. If desired, spark can be retarded and/or fuel may be reducedso that engine torque doesn't increase as the engine volumetricefficiency is increased. Thus, the cylinder flow and exhaust temperatureare increased and deliver more energy to accelerate the turbochargerturbine. If the routine determines that cylinder flow compensation orsurge control is desired, the routine proceeds to step 211. Otherwise,the routine proceeds to exit.

In step 211, adjustments are made to vary compressor flow.

In one embodiment, the relationship between the intake valve opening(IVO) position and engine volumetric efficiency is used to determinevalve adjustments that vary cylinder flow for both cylinder banks of aengine. Further, cylinder flow rates are used to adjust the downstreamturbine flow rates, thereby adjusting the flow of two compressors. Inparticular, the present position of each cam, or valve opening timingfor each cylinder bank, is evaluated with respect to the location ofbest volumetric efficiency to determine how to adjust the valve timing.Then, one cam is advanced while the other cam is retarded. This causesflow from one compressor to increase while flow from the othercompressor is reduced. The valve timing adjustments also cause theoverall cylinder air flow rate to remain substantially constant whilethe compressor flow rates are converging to an equalized flow rate. Forexample, if best volumetric efficiency occurs when both IVO and EVCoccur at top-dead-center (TDC) and if intake valves of both cylinderbanks of a “V” engine open 10 degrees before TDC (i.e., it is advanced)and a reduction in cylinder flow is desired in one cylinder bank, thenthe valve timing of the cylinder bank where lower flow is desired can befurther advanced (e.g., 15 degrees before TDC) to reduce the cylinder orcylinder bank mass flow rate. On the other hand, the valve timing of theother cylinder bank can be simultaneously retarded, therebycounteracting the valve timing advance on the opposite cylinder bank.Additional details describing the relationship between volumetricefficiency, IVO, and exhaust valve closing (EVC) are given in thedescription of FIG. 3.

Desired. IVO locations can be determined using the following equations:ivo_des_(—)1=ivo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))+K_(c)*comp_errorivo_des_(—)2=ivo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))K_(c)*comp_errorwhere ivo_des_1 is the desired IVO timing for cylinder bank number one,ivo_des(N,Γ_(eng) _(—) _(brake) _(—) _(tor)) is the base IVO timingdetermined from step 207 above, and K, is a predetermined gain that maytake on one or more values. For example, K_(c) may be a constant or itmay be a variable that is determined by indexing a table using enginespeed and load.

In another embodiment, cylinder flow can be adjusted by changing theposition of a waste gate and by adjusting valve timing. In engineoperating regions where only the waste gate is inactive, the followingmay be used to determine actuator commands:

If P_boost_des<P*_boost ivo_des_1=ivo_des(N,Γ_(eng) _(—) _(brake) _(—)_(tor))+K_(c)*comp_error ivo_des_2=ivo_des(N,Γ_(eng) _(—) _(brake) _(—)_(tor))−K_(c)*comp_error Else ivo_des_1=ivo_des(N,Γ_(eng) _(—) _(brake)_(—) _(tor)) ivo_des_2=ivo_des(N,Γ_(eng) _(—) _(brake) _(—) _(tor))WG_des_1=WG_des(N,Γ_(eng) _(—) _(brake) _(—)_(tor),P_boost)+K_(w)*comp_error WG_des_2=WG_des(N,Γ_(eng) _(—) _(brake)_(—) _(tor),P_boost)−K_(w)*comp_errorwhere WG_des_1 and WG_des_2 are the desired waste gate positioncommands, WG_des is the desired waste gate position, P_boost_des is thedesired boost pressure, P*_boost is the boost pressure at which thewastegate actuator achieves authority, and K_(w) is a waste gate gainterm which may be a constant or variable. When both actuators may beused simultaneously, intake valve timing and waste gate adjustment canbe accomplished with the following equations:ivo_des_(—)1=ivo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))+K_(c)*comp_errorivo_des_(—)2−ivo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))−K_(c)*comp_errorWG_des_(—)1=WG_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor) ,P_boost)+K_(w)*comp_errorWG_des_(—)2=WG_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor) ,P_boost)−K_(w)*comp_error

The above equations allow cylinder mass flow adjustments to be made toequalize flow between compressors during varying operating conditions.In one example, compressor flow can be equalized by reducing the massflowing through a cylinder bank to lower flow through a firstturbocharger turbine while the mass flow of the other cylinder bank isincreased through a second turbocharger turbine. In another example, themass flowing through two cylinder banks can be reduced to decrease flowthrough two turbocharger turbines and prevent compressor surge. In stillanother example, mass flowing through two cylinder banks can beincreased improve flow through two turbocharger turbines, therebyincreasing the compressor flow rate.

Specific intake and exhaust valve timing adjustments may be empiricallydetermined and stored in memory when a cylinder bank flow adjustment isrequested. Or alternatively, valve timings can be revised by findinggeometric solutions that are used to adjust engine exhaust residuals(i.e., combusted air-fuel mixtures) and cylinder air charge to desiredamounts. In particular, the method described in U.S. Pat. No. 6,754,575,which is hereby fully incorporated by reference for all, purposes, maybe used to determine valve timings. Further, waste gate or vanepositions can be scheduled and varied as intake and exhaust valve timingis varied by simply adding an exhaust valve adjustment term that issimilar to the intake valve adjustment term ivo_des_1.

Similarly, where valve lift is adjusted, a multi-dimensional tableincluding engine speed, intake manifold pressure, and valve lift amountcan be used in conjunction with compressor flow error to determine valvelift adjustments that will adjust turbine flow by adjusting the cylinderflow rate.

In addition, the throttle position determined in step 203 can be updatedso that the engine delivers the desired torque even after valve timing,waste gate, and/or lift adjustments have been made. For example, if avalve timing adjustment is made to increase the cylinder flow rate whilethe engine torque demand remains constant, the throttle position can beadjusted by an increment that represents the corresponding flow ratereduction that is desired in the opposite cylinder bank to balance thetorque between the cylinder banks. Thus, in this example, the throttleadjustment counteracts the additional torque that can result from thevalve adjustment.

In step 213, the desired valve operational adjustments are output toactuators. In one example, the camshaft angular position with respect tocrankshaft timing is converted into duty cycle signals that cause thecam actuators to advance or retard through the control of oil flow tothe cam phase actuator. The camshaft phase angle controller may simplyoutput commands from steps 207 and 211 that position the camshaft in anopen loop manner, or the controller may utilize camshaft positionfeedback and a proportional/integral position controller. Valve liftamounts are similarly processed. In system configurations that utilizeother types of variable valve actuators, electrically or hydraulicallycontrolled actuators for example, timing or lift adjustment can besimply be made to revise the timing that driver circuitry actuatesparticular valve actuators. The routing proceeds to step 215.

In step 215, the desired fuel mass charge is determined. The compensatedcylinder air charge mass is multiplied by the desired air-fuel ratio todetermine the desired fuel mass. The routine exits after the fueldelivery has been updated.

Referring now to FIG. 26, a flow chart of a method for simultaneouslycontrolling an engine cylinder mixture and regulating operatingconditions of a turbocharger is shown.

The method of FIG. 26 controls cylinder mixtures of a turbochargedengine while simultaneously controlling turbocharger operation. Themethod describes using different types of actuators to balance residualconcentrations in cylinders of a two cylinder bank engine. The methodmay be applied to an engine having one or more turbochargers. In oneembodiment, valve timing and/or lift can be used to regulate cylindermixtures while exhaust backpressure and turbine flow are varied byadjusting the position of a turbocharger waste gate. In another example,valve timing and/or lift adjustments can be combined with throttle plateposition adjustments and waste gate adjustments to control cylinderresiduals and cylinder air charge. Further, timing adjustments may bemade on per cylinder or per cylinder bank basis. That is, valveoperations for an individual cylinder or all cylinders of a cylinderbank may be made. This method can improve engine emissions and enginetorque, at least under some conditions.

In another aspect of the present system and method, the cylinder mixturecontents of individual cylinders can be adjusted such that cylindermixture regulation is not limited to banks or groups of cylinders. Thatis, adjustments may be directed to individual cylinders so that cylindermixture variations between individual cylinders are reduced. This may bebeneficial when there are differences in flow paths leading to and fromdifferent cylinders. In one example, individual cylinder adjustments canbe accomplished by electrically actuated valves. These valves can beoperated such that each cylinder has valve timing that is different fromother cylinders, if desired.

Steps 250-254 of FIG. 2B are identical to steps 201-205 in FIG. 2A. Assuch, the description of steps 201-205 apply to steps 250-254, thedescription for steps 250-254 is omitted to make a more concisedescription.

In step 256, the routine determines the desired residual mass andestimates the actual residual mass. The desired residual mass isdetermined by interpolating between cells of tables or functions thatcontain empirically determined data that describe desirable amounts orpercentages of residuals for selected engine operating conditions. Thetables or functions are indexed based on engine speed and engine load.Typically, both cylinder banks operate with the same cylinder residualamount but it is possible to have different residual amounts for eachcylinder bank.

The actual cylinder residual amount for each cylinder bank can beestimated by interrogating a set of tables or function similar to thoseshown in FIGS. 5 and 6. The functions are empirically determined anddescribe the relationship between the position of a cam or valve timing,cylinder airflow, and intake manifold pressure. The tables or functionsdescribe representations similar to those illustrated in FIGS. 5 and 6.Further, these representations can be compensated for variations incylinder back pressure by adjusting the offset term that is describedbelow, c₂·(P_(exh)/P_(nom)), in relation to the exhaust backpressure.The relationships illustrated in the lines shown in FIGS. 5 and 6 can beexpressed as the equation of a straight line taking the form:

$P_{m} = {{c_{1} \cdot m_{air}} + {c_{2} \cdot \frac{P_{exh}}{P_{nom}}}}$where the intake manifold pressure, m_(air) is the mass of air in thecylinder, P_(exh) is the exhaust manifold pressure at the currentoperating condition, c_(a) and c₂ are functions of intake and exhaustvalve timing or cam position, and P_(nom) is the manifold pressure withthe engine off (barometric pressure). Rearranging to solve for m_(air)gives:

$m_{air} = {\left( \frac{P_{m}}{c_{1}} \right) - {\left( {\frac{c_{2}}{c_{1}} \cdot \frac{P_{exh}}{P_{nom}}} \right).}}$Thus, the residual mass, r_(m), is defined as:

$r_{m} = \left( {\frac{c_{2}}{c_{1}} \cdot \frac{P_{exh}}{P_{nom}}} \right)$The residual mass is determined by indexing equations of lines that aresimilar to those represented in FIGS. 5 and 6 using the present campositions or valve timings, engine load, and engine speed. These dataare then used to determine the current residual mass as described above.The routine proceeds to step 258.

Two versions of steps 258-262 are presented to describe two embodiments.The first embodiment is for an engine having two cylinder banks and asingle turbocharger. The second embodiment is for an engine having twocylinder banks and two turbochargers, one turbocharger per cylinderbank.

Single turbocharger embodiment: In step 258, the open-loop camshaftpositions or valve timings for each cylinder bank are determined.Alternatively, valve lifts may be determined in this step insubstantially the same manner, i.e., by indexing predetermined tablesand/or functions. Engine speed and desired engine torque are used toindex tables that output empirically predetermined cam positions orvalve timings for each cylinder bank. In this embodiment, the open-loopvalve timings are set so that overlap between intake and exhaust valvesis increased as engine speed increases for the cylinder bank that doesnot have a turbocharger in its exhaust path. This increases thecylinder's residual mass fraction to account for the increased cylinderback pressure of the cylinder bank having a turbocharger in its exhaustpath. The cams positions or valve timings can be adjusted such thatvariation of cylinder mixtures between cylinder banks is reduced. Theparticular cam timings or valve timings used to reduce cylinder mixturevariation can be empirically determined at steady state engine operatingconditions during engine mapping. The routine proceeds to step 260.

In step 260, the routine determines whether or not to make valve timing,valve lift, throttle, and turbocharger adjustments (e.g., repositioningthe waste gate or turbine vanes) to correct for compressor flowdifferences and/or residual differences between cylinder banks. Todetermine whether or not valve adjustments are desired, the estimatedresiduals between each cylinder bank are compared. Cylinder residualcomparisons are made as follows:bank1_res_error=mres_(des)−bank1_(—) m _(d)bank2_res_error=mres_(des)−bank2_(—) m _(d)where bank1_res_error is the residual error for cylinder bank numberone, mres_(des) is the desired residual mass determined in step 256,bank1_m_(d) is cylinder bank number one estimate of residual,bank2_res_error is the residual error for cylinder bank number two, andbank2_m_(d) is cylinder bank number two estimate of residual. The termsbank1_m_(d) and bank2_m_(d) may be determined as described in step 256.Alternatively, the residual errors may be determined by the followingequation:res_error=bank1_(—) m _(d)−bank2_(—) m _(d)where res_error is the residual error between the two cylinder banks,and bank1_m_(d) and bank2_m_(d) are as described above. The res_errormay be substituted for the bank1_res_error and bank2_res_error terms inthe following equations when signs of the variables are taken intoaccount.

After the cylinder bank exhaust residual mass errors are determined, theroutine determines if the errors are large enough to warrant correctionsto valve timing and lift. If desired, the routine can limit residualcorrections to one or more predetermined magnitudes of error. That is,if the residual error between the two cylinder banks is below somepredetermined value, then no correction will be made to compensate forthe error. In another embodiment, the residual error between cylinderbanks can be compensated as long as there is an error present.

Continuing with step 260, the routine also determines if adjustments tovalve timing/lift, throttle position, and the turbocharger are to bemade to limit the possibility of compressor surge. If the turbochargeris operating near or goes into the compressor surge region, the valvetiming/lift can be adjusted such that a cylinder bank volumetricefficiency is decreased, thereby lowering the pressure ratio across thecompressor and reducing the possibility of compressor surge. Severaldifferent approaches can be used to decide whether or not to takemitigating steps to control compressor surge. In one embodiment, thecompressor surge boundary is stored in the engine controller memory. Ifthe compressor is operating within a predetermined amount of compressorflow then valve timing can be adjusted to reduce cylinder flow andmitigate the possibility of entering a compressor surge condition.Likewise, if the compressor is operating within a predetermined pressureratio amount of the compressor surge boundary, valve timing can be usedto reduce cylinder flow. If the routine determines that cylinderresidual compensation or compressor surge control is desired, theroutine proceeds to step 262. Otherwise, the routine proceeds to exit.

In step 262, adjustments are made to vary compressor flow and/orcylinder residuals.

In one embodiment, the cam timing of the cylinder bank that does nothave a turbine in its exhaust path is adjusted to match the cylinderresiduals between cylinder banks. Intake valve timing, exhaust valvetiming, or a combination of intake valve timing and exhaust valve timingadjustments may be used to accomplish balancing residual masses betweencylinder banks. In one example, where engine flow is increasing and theexhaust pressure is increasing on the cylinder bank having a turbine inits exhaust path, the exhaust cam is retarded on the opposite cylinderbank to increase the cylinder residuals. In this example, retarding theexhaust cam increases the amount of valve overlap between intake andexhaust valves. That is, the amount if time during the period whenintake and exhaust valves are simultaneously open is increased. Theexhaust cam retard may follow a predetermined table or function based onempirical testing, or alternatively, the rate of cam retard may be basedon the cylinder bank flow rate or on the pressure in the exhaustmanifold of the cylinder bank without a turbine in its exhaust path.

In another embodiment, the exhaust valve timing of both cylinder banksmay be adjusted to equalize cylinder residuals between the cylinderbanks. The exhaust valve timings can be described asevo_des_(—)1=evo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))+K_(e)*bank1_res_errorevo_des_(—)2=evo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))−K_(e)*bank2_res_errorwhere evo_des_1 is the exhaust valve desired opening position, evo_desis the open-loop desired exhaust valve opening position, and Ke is again term. Thus, the exhaust valve opening positions are varied ascylinder residuals vary from the desired amount. This control actioncauses the residuals between the cylinder banks to converge to the sameamount even though one cylinder bank has a turbine in its exhaust path.Adjusting the position of an exhaust cam causes the point at which linesdescribing cylinder air charge, as related to cam position and airflow,intersect zero air flow. This intersection identifies the amount ofcylinder residuals and therefore indicates that the amount of cylinderresidual exhaust gases change as a function of exhaust cam position.This concept can be seen from the separation of the lines in FIG. 5.

In addition, the throttle position determined in step 252 can be updatedso that the engine delivers the desired torque even after valveoperation and turbocharger adjustments have been made. For example, whena valve timing adjustment shifts the pumping characteristics of acylinder bank the throttle position is adjusted to compensate ofvariation of the cylinder mass flow rate that may result from theadjusted valve operation. The throttle position is updated by the samemethod presented in step 203. The routine proceeds to step 264.

Two turbocharger embodiment: In step 258, the open-loop camshaftpositions or valve timings for each cylinder bank are determined.Alternatively, valve lifts may be determined in this step insubstantially the same manner, i.e., by indexing predetermined tablesand/or functions. Engine speed and desired engine torque are used toindex tables that output empirically predetermined cam positions orvalve timings for each cylinder bank. Typically, the open-loop valveoperations (i.e., timing/lift) are set equally between the cylinderbanks, but it is possible to request different operations between thecylinder banks if desired. The routine proceeds to step 260.

In step 260, the routine determines whether or not to make valve timing,valve lift, throttle, and turbocharger adjustments to correct forcompressor flow differences and/or residual differences between cylinderbanks. As described in FIG. 2A, it can be desirable to substantiallyequalize flow rates between two compressors operating with two differentcylinder banks. Flow rate balancing can be accomplished by waste gateand/or valve operating adjustments. However, as described in thebackground, equalizing compressor flow can sometimes result inbackpressure differences between cylinder banks. If the residual gasfraction is different between the two cylinder banks valve operation canbe adjusted to compensate for this condition, at least during someengine operating conditions.

To determine whether or not valve adjustments are desired, a comparisonis made between the mass of air flowing through each compressor.Compressor flow errors are determined as described in step 209 of FIG.2A. Further, the estimated residuals between each cylinder bank arecompared. Cylinder residual comparisons are made as follows:bank1_res_error=mres_(des)−bank1_(—) m _(d)bank2_res_error=mres_(des)−bank2_(—) m _(d)where bank1_res_error is the residual error for cylinder bank numberone, mres_(des) is the desired residual mass determined in step 256,bank1_m_(d) is cylinder bank number one estimate of residual,bank2_res_error is the residual error for cylinder bank number two, andbank2_m_(d) is cylinder bank number two estimate of residual. The termsbank1_m_(d) and bank2_m_(d) may be determined as described in step 256.Alternatively, the residual errors may be determined by the followingequation:res_error=bank1_(—) m _(d)−bank2_(—) m _(d)where res_error is the residual error between the two cylinder banks,and bank1_m_(d) and bank2_m_(d) are as described above. The res_errormay be substituted for the bank1_res_error and bank2_res_error terms inthe following equations when signs of the variables are taken intoaccount.

After the cylinder bank exhaust residual mass errors are determined, theroutine can determines if the errors are large enough to warrantcorrections to valve timing and lift. If desired, the routine can limitresidual corrections to one or more predetermined magnitudes of error.That is, if the residual error between the two cylinder banks is belowsome predetermined value, then no correction will be made to compensatefor the error. In another embodiment, the residual error betweencylinder banks can be compensated as long as there is an error present.

Referring again to step 260, the routine also determines if adjustmentsto valve timing/lift, throttle position, and the turbocharger are to bemade to limit the possibility of compressor surge. If the turbochargeris operating near or goes into the compressor surge region, the valvetiming/lift can be adjusted such that a cylinder bank volumetricefficiency is decreased, thereby lowering the pressure ratio across thecompressor and reducing the possibility of compressor surge. Severaldifferent approaches can be used to decide whether or not to takemitigating steps to control compressor surge. In one embodiment, thecompressor surge boundary is stored in the engine controller memory. Ifthe compressor is operating within a predetermined amount of compressorflow then valve timing can be adjusted to reduce cylinder flow andmitigate the possibility of entering a compressor surge condition.Likewise, if the compressor is operating within a predetermined pressureratio amount of the compressor surge boundary, valve timing can be usedto reduce cylinder flow. If the routine determines that cylinderresidual compensation or compressor surge control is desired, theroutine proceeds to step 262. Otherwise, the routine proceeds to exit.

In step 262, adjustments are made to vary turbocharger flow and/orcylinder residuals.

During some engine operating conditions there are some systemconfigurations that have limitations that make it difficult tosimultaneously control turbocharger compressor flow and cylinderresidual mass by valve timing alone. For example, systems that areconfigured to have fixed intake valve timing and variable exhaust valvetiming with fixed valve lift duration. In this configuration, theopening and closing of exhaust valves can be varied relative to thecrankshaft. When the exhaust cam timing is adjusted to vary cylinderresiduals, the exhaust valve opening and closing positions move relativeto the crankshaft. Since exhaust valve closing position can also alterflow from the cylinder to the turbine, the rate of flow from a cylinderto the turbine may change in an undesirable way.

Recognizing this, the inventors herein have determined that it can bedesirable to simultaneously control a turbocharger using its waste gateand to control cylinder residuals using engine valve operatingadjustments. When the waste gate and engine valves are controlledtogether, the adjustments to the waste gate and valves can be made toaccount for changes to the operation of one actuator that affect thecontrol of the other actuator. For example, when the waste gate positionis adjusted to vary boost pressure, valve operation can be adjusted tocompensate for pressure changes in the exhaust system. Thus, theturbocharger waste gate and valves are adjusted together, and eachactuator compensates for the effects of adjusting the other actuator, atleast to some extent.

In one embodiment, where exhaust valve timing is adjustable, exhaustvalve timings are used to vary cylinder residuals while turbochargerwaste gate positions are adjusted to account for compressor flowimbalances. The waste gate commands are as described in step 213 of FIG.2A:WG_des_(—)1=WG_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor) ,P_boost)+K_(w)*comp_errorWG_des_(—)2=WG_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor) ,P_boost)−K_(w)*comp_error

The exhaust cam timings areevo_des_(—)1=evo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))+K_(e)*bank1_res_errorevo_des_(—)2=evo_des(N,Γ _(eng) _(—) _(brake) _(—) _(tor))−K_(e)*bank2_res_errorwhere evo_des_1 is the exhaust valve desired opening position, evo_desis the open-loop desired exhaust valve opening position, and Ke is again term. Thus, the exhaust valve opening positions are varied ascylinder residuals vary from the desired amount. The result of thisaction is that the cylinder exhaust residuals change as the exhaustvalve opening position is adjusted. Specifically, adjusting the positionof an exhaust cam causes the point at which lines describing cylinderair charge, as related to cam position and airflow, intersect zero airflow. This intersection identifies the amount of cylinder residuals andtherefore indicates that the amount of cylinder residual exhaust gaseschange as a function of exhaust cam position. This concept can be seenfrom the separation of the lines in FIG. 5.

In addition, the throttle position determined in step 252 can be updatedso that the engine delivers the desired torque even after valveoperation and turbocharger adjustments have been made. For example, whena valve timing adjustment shifts the pumping characteristics of acylinder bank the throttle position is adjusted to compensate ofvariation of the cylinder mass flow rate that may result from theadjusted valve operation. The throttle position is updated by the samemethod presented in step 203. The routine proceeds to step 264.

In step 264, the desired valve timings or lift amounts are output toactuators. In one example, the camshaft angular position with respect tocrankshaft timing is converted into duty cycle signals that cause thecam actuators to advance or retard through the control of oil flow tothe cam phase actuator. The camshaft phase angle controller may simplyoutput commands from steps 258 and 264 that position the camshaft in anopen loop manner, or the controller may utilize camshaft positionfeedback and a proportional/integral position controller. Valve liftamounts are similarly processed. The routine proceeds to step 266.

In step 266, the desired fuel mass charge is determined. The compensatedcylinder air charge mass is multiplied by the desired air-fuel ratio todetermine the desired fuel mass. The routine exits after the fueldelivery has been updated.

Note that turbine vane position adjustments can be readily substitutedfor waste gate adjustments in the above description, if desired. Alsonote that spark adjustments may be included in the methods described inFIGS. 2A and 2B. Spark can be adjusted as a function of cylinder load,engine speed, and residuals. Additionally, spark may be adjusted toincrease or decrease the amount of energy being delivered to a turbine.Further, the methods described in FIGS. 2A and 2B can includeadjustments for the fueling amount and timing. Likewise, fueladjustments can be used to increase or decrease the amount of energydelivered to a turbine. Further still, timing adjustments to valveoperation are made in relation to crankshaft position. That is, if anexhaust cam is retarded 5 degrees, the 5 degrees are measured as 5degrees of crankshaft, rotation.

Referring now to FIG. 3, a plot of a turbocharger compressor map isshown. The Y-axis describes the pressure ratio across the turbochargercompressor. That is, the ratio created by dividing the compressor outputpressure by the compressor inlet pressure. The X-axis describes thecompressor flow rate. Curves 301-305 represent pressure ratio versusflow for different compressor speeds. Curve 310 can be stored in enginecontroller 12 memory and represents the compressor surge boundary whereflow through the compressor degrades. Accordingly, operating thecompressor in the region to the left of curve 310 is undesirable. If thecompressor pressure ratio is driven toward the surge line, by a rapidlychanging throttle plate position for example, the turbocharger's turbinewaste gate or vane position can be adjusted along with valve timing/liftto mitigate the effects of operating in this region. In this way, thetime that turbocharger operates in an undesirable condition may bereduced. In one embodiment, exhaust valve timing can be retarded tolower the volumetric efficiency of a cylinder bank and reduce cylinderfollow to a turbine that is coupled to a compressor that is approachinga surge boundary. In another embodiment, intake valve timing of acylinder bank can be retarded to reduce cylinder flow to a turbine thatis coupled to a compressor that is approaching a surge boundary. Instill another embodiment, intake valve timing of a cylinder bank can beadvanced to reduce cylinder flow to a turbine that is coupled to acompressor that is approaching a surge boundary.

Referring now to FIG. 4, a plot that describes the relationship betweenexample engine valve timing and engine volumetric efficiency is shown. Asurface is created for illustration purposes by linking together mappedengine operating points. The Y-axis represents engine volumetricefficiency (VOL EFF). Volumetric efficiency is defined here as the ratioof actual air pumped through the engine to the theoretical engine airpumping capacity. Efficiency is normalized between zero and one, whereone is equal to 100% efficiency. The X-axis represents the IVO location(or timing) in degrees relative to TDC intake stroke (i.e., zerodegrees). Numbers less than zero represent intake valve openinglocations that are before TDC and numbers greater than zero representintake valve opening locations that are after TDC. The Z-axis representsexhaust valve closing (EVC) location (or timing) in degrees relative toTDC. In this example, all exhaust valve closing times are shown retardedfrom TDC; however, exhaust valve closing times may be advanced from TDCif desired.

In curve 401, for example, volumetric efficiency decreases (i.e., lessair is pumped through the engine) when IVO is advanced (moved to theright of TDC) or retarded from TDC. This curve represents the influenceof IVO on engine volumetric efficiency when EVC is held constant atforty degrees. Advancing or retarding valve timings for other operatingconditions similarly influences volumetric: efficiency, except that thelocation of higher volumetric (i.e., TDC in FIG. 401) varies with engineoperating conditions.

The relationship between EVC and volumetric efficiency is illustrated bycurve 405. Here, volumetric efficiency decreases as EVC is retarded fromTDC.

Several observations can be made from the surface plot. Namely,depending on the initial valve timings, advancing intake valve timing atone condition will reduce engine volumetric efficiency while advancingvalve timing from another condition will increase engine volumetricefficiency. Therefore, the decision of whether to advance or retardintake valve timing will depend on the initial intake valve timing andthe desire to increase or decrease cylinder volumetric efficiency. Thefigure also shows that exhaust valve timing can be retarded to reduceengine volumetric efficiency and cylinder flow. By retarding the exhaustvalve timing, more residual gas is trapped in the cylinder and thepropensity for fresh air to be inducted into the cylinder is reduced.

Referring now to FIG. 5, a plot that illustrates the effect of variableexhaust valve timing on intake manifold pressure and cylinder air chargeis shown. The data were generated from a naturally aspirated engineoperating at 1500 RPM but turbocharged engines are expected to exhibitsimilar characteristics. The Y-axis represents intake manifold absolutepressure (MAP) in units of inches of mercury. The X-axis represents theengine airflow in grams per second. Line 501 represents the relationshipbetween intake manifold pressure and engine airflow when the intakevalve opens at 10 degrees before top-dead-center intake stroke and theexhaust valve closes at 40 degrees after top-dead-center intake stroke.Line 503 represents the relationship between intake manifold pressureand engine airflow when the intake valve opens at 10 degrees beforetop-dead-center intake stroke and the exhaust valve closes at 30 degreesafter top-dead-center intake stroke. Line 505 represents therelationship between intake manifold pressure and engine airflow whenthe intake valve opens at 10 degrees before top-dead-center intakestroke and the exhaust valve closes at 20 degrees after top-dead-centerintake stroke. Line 507 represents the relationship between intakemanifold pressure and engine airflow when the intake valve opens at 10degrees before top-dead-center intake stroke and the exhaust valvecloses at top-dead-center intake stroke. Thus, the intake manifoldpressure and full load cylinder air charge can be affected by the choiceof exhaust valve timing. The figure also illustrates that intakemanifold pressure can be readily determined given an engine airflow anda cam position, as is described in the explanation of FIG. 2.

Referring now to FIG. 6, a plot that illustrates the effect of variableintake valve timing on intake manifold pressure and cylinder air chargeis shown. The data were generated at 1500 RPM. The Y-axis representsintake manifold absolute pressure (MAP) in units of inches of mercury.The X-axis represents the engine airflow in grams per second. Line 621represents the relationship between intake manifold pressure and engineairflow when the intake valve opens at 30 degrees after top-dead-centerintake stroke and the exhaust valve closes at 20 degrees aftertop-dead-center intake stroke. Line 623 represents the relationshipbetween intake manifold pressure and engine airflow when the intakevalve opens at 20 degrees after top-dead-center intake stroke and theexhaust valve closes at 20 degrees after top-dead-center intake stroke.Line 625 represents the relationship between intake manifold pressureand engine airflow when the intake valve opens at 30 degrees beforetop-dead-center intake stroke and the exhaust valve closes at 20 degreesafter top-dead-center intake stroke. Line 627 represents therelationship between intake manifold pressure and engine airflow whenthe intake valve opens at top-dead-center intake stroke and the exhaustvalve closes at 20 degrees after top-dead-center intake stroke. Theseveral curves illustrate that a desired manifold pressure can bereadily determined at different cam positions and engine speeds. Thisrelationship can be used to advantage as is described in the explanationof FIG. 2.

As will be appreciated by one of ordinary skill in the art, the routinesdescribed in FIG. 2 may represent one or more of any number ofprocessing strategies such as event-driven, interrupt-driven,multi-tasking, multi-threading, and the like. As such, various steps orfunctions illustrated may be performed in the sequence illustrated, inparallel, or in some cases omitted. Likewise, the order of processing isnot necessarily required to achieve the objects, features, andadvantages described herein, but it is provided for ease of illustrationand description. Although not explicitly illustrated, one of ordinaryskill in the art will recognize that one or more of the illustratedsteps or functions may be repeatedly performed depending on theparticular strategy being used.

This concludes the description. The reading of it by those skilled inthe art would bring to mind many alterations and modifications withoutdeparting from the spirit and the scope of the description. For example,I3, I4, I5, V6, V8, V10, and V12 engines operating in natural gas,gasoline, diesel, or alternative fuel configurations could use thepresent description to advantage.

The invention claimed is:
 1. A method for controlling a V-engine withfirst and second cylinder banks, comprising: during exhaust backpressuredifferences between the banks, varying valve operation of cylinders inthe second bank to increase torque as a waste gate position of a firstturbocharger is adjusted, said first turbocharger having a turbinelocated in an exhaust of the first bank and a second turbocharger havinga turbine located in an exhaust of the second bank; while varyingposition of a throttle to reduce flow rate and counteract the torqueincrease.
 2. The method of claim 1 further comprising adjusting spark insaid second bank of cylinders as said valve operation is varied in saidsecond bank.
 3. The method of claim 1 wherein varying valve operationincludes varying intake valve timing relative to a crankshaft of theengine.
 4. The method of claim 1 wherein varying valve operationincludes varying exhaust valve timing relative to a crankshaft of theengine.
 5. A method for controlling an engine, comprising: balancingmass flow through first and second engine cylinder groups varying valveoperation of cylinders in the second cylinder group as a waste gateposition of a first turbocharger is adjusted and while advancingignition spark, said first turbocharger located in an exhaust path ofthe first cylinder group, the mass flow through the second cylindergroup directed to a second turbocharger.
 6. The method of claim 5further comprising varying valve operation of cylinders in the firstcylinder group oppositely to adjusting of the cylinders in the secondcylinder group.
 7. The method of claim 6 wherein the varying valveoperation of cylinders in the first cylinder group and the varying valveoperation of cylinders in the second cylinder group comprises advancinga cam timing of the first cylinder group cylinders to advance valveopening timing while retarding a cam timing of the second cylinder groupcylinders to retard valve opening timing, the flow balanced byincreasing flow of a compressor of the first turbocharger whiledecreasing flow of the second turbocharger and while maintaining anoverall cylinder air flow substantially constant.
 8. The method of claim5 wherein a waste gate of said second turbocharger is in communicationwith said second cylinder group.
 9. The method of claim 5 wherein thevarying valve operation of cylinders in the second cylinder group variesvalve timing relative to a crankshaft.
 10. The method of claim 5 whereinsaid valve operation is varied as an amount of exhaust residuals vary insaid first cylinder group.
 11. The method of claim 5 wherein said valveoperation is varied by varying valve lift.
 12. The method of claim 5wherein varying said valve operation increases an amount of valveoverlap between intake and exhaust valves of said cylinders.
 13. Themethod of claim 5 further comprising adjusting a position of a throttleplate as said valve operation is varied.
 14. The method of claim 5wherein said valve operation is a timing of an exhaust valve relative toa crankshaft position.
 15. The method of claim 5 wherein said valveoperation is a timing of an intake valve relative to a crankshaftposition.
 16. A method for controlling an engine, comprising: balancingmass flow through first and second engine cylinder groups varying valvetiming operation of cylinders in the second cylinder group as a wastegate position of a first turbocharger is adjusted, said firstturbocharger located in an exhaust path of the first cylinder group, themass flow through the second cylinder group directed to a secondturbocharger; while varying a throttle position upstream of bothcylinder groups.
 17. The method of claim 16 further comprising varyingvalve operation of cylinders in the first cylinder group oppositely toadjusting of the cylinders in the second cylinder group.
 18. The methodof claim 17 wherein the varying valve operation of cylinders in thefirst cylinder group and the varying valve operation of cylinders in thesecond cylinder group comprises advancing a cam timing of the firstcylinder group cylinders to advance valve opening timing while retardinga cam timing of the second cylinder group cylinders to retard valveopening timing, the mass flow balanced by increasing flow of acompressor of the first turbocharger while decreasing flow of the secondturbocharger and while maintaining an overall cylinder air flowsubstantially constant.
 19. The method of claim 16 wherein a waste gateof said second turbocharger is in communication with said secondcylinder group.
 20. The method of claim 16 wherein the varying valveoperation of cylinders in the second cylinder group varies the valvetiming relative to a crankshaft.